Torsional tunable coupling for a diesel engine drive shaft

ABSTRACT

An improved coupling assembly is provided for transmitting rotary power from the working end of an internal combustion engine to a driven shaft. The crankshaft also has a free end connected to an accessory drive train. The coupling assembly comprises a low inertia flywheel having a mass so selected as to cause the node of the first crankshaft mode of torsional vibration to be located in the vicinity of the middle of the crankshaft, and a flexible coupling which interconnects the working end of the crankshaft with the driven shaft. The low inertia flywheel not only reduces the amplitude of the torsional deflection at the free end of the crankshaft, but further raises the primary torsional vibration orders of the engine which excite the coupling assembly by at least one half of an order such that the peak stresses applied to the teeth of the first gear wheel of the accessory drive train are at least halved and are further applied to at least twice as many gear teeth, thereby greatly prolonging the life of the gear wheel. Moreover, the low inertia flywheel further increases the lifetime of the flexible coupling by raising its natural frequency to a level which is substantially higher than the 0.5 engine order of torsional vibration associated with engine malfunction and governor interaction.

This application is a continuation of Ser. No. 08/228,759, filed Apr.18, 1994, now abandoned; which is a continuation of Ser. No. 08/032,414,filed Mar. 15, 1993, now U.S. Pat. No. 5,303,681, which is acontinuation of Ser. No. 07/936,524, filed Aug. 28, 1992, now abandoned.

BACKGROUND OF THE INVENTION

This invention generally relates to engine couplings, and isspecifically concerned with an improved, low-inertia coupling assemblyfor reducing stress in the interface between the ends of a crankshaftand a drive train.

Coupling assemblies for transmitting rotary power from the working endof a crankshaft of an internal combustion engine to a driven shaft arewell known in the prior art. Such coupling assemblies generally comprisea high inertia flywheel in combination with a flexible coupling whichinterconnects the working end of the crankshaft with the driven shaft.The flexible coupling may include a resilient member formed from anelastomeric material. Such couplings are most typically used in dieselengines, and the primary purpose of the high-inertia flywheel is tosmooth out the amplitude of the torque generated by the working end ofthe crankshaft. A secondary purpose of the flywheel is to provide amount for the ring gear which engages the output gear of the startermotor of the engine. The flexible coupling utilized in such prior artassemblies not only serves the function of mechanically interconnectingthe working end of the crankshaft with a driven shaft; the flexibilityprovided by the elastomeric material in the coupling advantageouslydampens impulse torques which might otherwise be generated between thecrankshaft and the driven shaft. Such unwanted impulse torques mayoccur, for example when the driven shaft is a cardan-type shaft, and theelastomeric material provided in such a flexible coupling allows thecoupling to drive such a shaft for a maximum amount of time withoutfailure.

While such prior art coupling assemblies have performed satisfactorilyin the past, the applicant has observed a number of shortcomings in theperformance of such couplings as the power of diesel engines hasincreased over the years. For example, the applicant has noted that therelatively large inertias associated with the flywheels of prior artcouplings (typically between 150 and 400 lbs*ft² in diesel engines ofbetween about 500 and 2000 horsepower) tend to cause the node of thefirst mode of crankshaft torsional vibration to be located in thevicinity of the flywheel itself. Such a location has the effect ofmaximizing the amplitude of the torsional vibration experienced by thefree end of the crankshaft. Since the free end of the crankshaft of suchdiesel engines is typically connected to an accessory drive train suchas the timing gear train and vehicle accessory drives, the relativelylarge amplitude of torsional movement of the free end of the crankshaftcreates undesirable stress in this gear train which is particularlyintense with respect to the teeth of a crank nose pinion of the geartrain.

The applicant has also observed three other major problems that comeabout as a result of the relatively large mass of the flywheels used insuch prior art couplings. The first and most important of these problemsis concentration of intense stress on only a few of the gear teeth ofthe gear train driven by the free end of the crankshaft. The inherentnatural frequency of the crankshaft mode (or second system mode) oftorsional vibration causes the crankshaft to be excited by relativelylow engine orders (such as the second, third or fourth orders in a four,six or eight throw diesel crank respectively). Hence, in the case of aneight throw crank, the excitation of the crankshaft mode of torsionalvibration in the engine of a prior art flywheel assembly by the fourthengine order results in the same four teeth (located 90° apart) beingsubjected to very high torsional vibrational stresses with eachrevolution of the gear wheel. After a period of time, these stressescause these four gear teeth to fail, thereby necessitating an expensiveand time-consuming replacement of the gear wheel. A second problemassociated with the use of a high inertia flywheel in such prior artcouplings is the relatively low frequency it confers on the couplingmode (first system mode); i.e. frequencies in the range of 15 to 20hertz. While these frequencies avoid major exciting orders in the engineoperating speed range, the engine has to run through the couplingresonance speed during start-up, and can damage the coupling byexcessive deflections at such low frequencies.

A third problem occurs if this low coupling mode frequency brings thehalf-order resonance speed within the upper speed range of the engine.Either a misfiring cylinder, or vigorous governor action will cause ahigh level of half-order excitation which can damage or break thecoupling under these conditions.

Other shortcomings associated with the use of such a high-inertiaflywheel include the out-of-balance and bending moment forces that sucha flywheel applies to the crankshaft which supports it, as well as theexpense necessitated by the precision manufacture and installation ofsuch heavy components in an engine.

Clearly, there is a need for an improved coupling assembly whichovercomes the shortcomings and problems associated with the use ofhigh-inertia flywheels in such assemblies.

SUMMARY OF THE INVENTION

The invention is both an apparatus and a method which eliminates or atleast ameliorates all the problems associated with prior art couplingassemblies that employ high-inertia flywheels which cause the node ofthe first crankshaft mode to be located in the vicinity of the flywheel.Specifically, the improved coupling assembly of the invention employs alow inertia flywheel having a mass so selected as to cause the node ofthe first crankshaft mode of torsional vibration to be located in thevicinity of the middle of the crankshaft, which not only reduces theamplitude of the torsional deflection at the free end of the crankshaft,but also changes the primary engine orders that excite the couplingassembly from whole number orders to half-orders. These two effectscombine to greatly reduce the stress at the interface between the freeend of the crankshaft and a drive train connected to this end. Forexample, where this interface is defined by the gear teeth of a timinggear, the changing of the primary exciting orders of torsional vibrationby at least one half of an order reduces the peak stresses applied tothe teeth of the gear by at least one half, and applies these stressesto at least twice as many gear teeth, thereby greatly prolonging thelifetime of the timing gear.

The improved coupling assembly of the invention preferably includes aflexible coupling which employs an elastomeric element for connectingthe working end of the crankshaft with a driven shaft. Referring to FIG.5, showing a bearing supported shaft, the arrangement of this componentof the coupling assembly increases the versatility of the assembly byallowing the driven shaft to be a cardan type shaft. The use of alow-inertia flywheel in the coupling assembly advantageously increasesthe natural frequency of the combination of the flexible coupling andthe flywheel to a level which is substantially higher than the 0.5engine order associated with governor action and engine malfunction,thereby eliminating or at least greatly diminishing the probability ofcoupling failure as a result of spurious excitation over the life of theengine. In the preferred embodiment, the flexible coupling is ashear-block coupling which can act as a mechanical "fuse" should therelative torque between the crankshaft and driven shaft exceed apredetermined safe level.

The coupling assembly of the invention may further include a means fortuning or adjusting the mode of the coupling assembly formed from amanually removable retaining ring for facilitating the replacement ofthe elastomeric member of the flexible coupling with another elastomericmember having different hardness characteristics.

The driven shaft may be a floating shaft, and the flexible coupling mayfurther include a centering ring for maintaining the concentricity ofone end of the floating shaft with the flexible coupling. Additionally,one edge of the centering ring may be connected to the retaining ringand the centering ring may circumscribe the driven member of theflexible coupling. The use of such a centering ring obviates the needfor one of the bearing assemblies which normally rotatably supports sucha floating shaft.

Alternatively, the coupling assembly of the invention may integrallyinclude a bearing assembly for rotatably supporting one end of thedriven shaft, which would not only obviate the need for a separatebearing assembly to be constructed somewhere along the length of thedriven shaft, but which would also conveniently ensure an on-center,concentric relationship between the coupling assembly and one end of thedriven shaft. A housing is also preferably provided that not onlyencloses both the flywheel and the flexible coupling, but also supportsthe previously mentioned, integrally-provided bearing assembly as well.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic representation of a prior art diesel engine systemin which the engine coupling assembly of the present invention can beemployed.

FIG. 2 is a detailed side view of one embodiment of the engine couplingassembly, depicting the present invention, in which the driven couplingis carried by a two bearing shaft.

FIG. 3 is a cross sectional view of the flexible coupling assembly ofFIG. 2 taken along line 3--3.

FIG. 4 is an alternate embodiment of the engine coupling assembly shownin FIG. 2 in which a centering ring is used to support a floating shaftdrive suitable for use where precise alignment is not needed and quickchanges are desirable.

FIG. 5 is an alternate embodiment of the engine coupling assembly shownin FIG. 2 in which an integral bearing assembly suitable to drive acardan shaft or single bearing generator is provided.

FIG. 6 is a Campbell diagram showing the relationship between systemmode frequencies, engine RPM, and engine excitation orders.

FIG. 7 is a graph illustrating the relationship between engines having4, 6 and 8 crank throws and the relative magnitudes of the variousorders of torsional vibration.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

FIG. 1 shows a prior art diesel engine system of which the currentinvention can be a part. Shown in FIG. 1 is crank pinion gear 118,diesel engine 110, engine coupling assembly 10, and output shaft 119.Diesel engine 110 is a four crank throw engine of the type includingfour pistons 112 connected through connecting rods 114 to enginecrankshaft 116. Engine crankshaft 116 has a free end 115 and a flywheelend 117. A four crank throw diesel engine is shown in FIG. 1, but anengine of any type with any number of crank throws and cylinders couldbe used with the present invention.

The free end 115 of engine crankshaft 116 is connected to crank piniongear 118. Typically, crank pinion gear 118 is used to drive a timinggear train and/or vehicle accessory drives (shown schematically inphantom). The flywheel end 117 of engine crankshaft 116 is connected toengine coupling assembly 10. Also connected to engine coupling assembly10 is output shaft 119.

In FIG. 1, engine coupling assembly 10 contains flywheel 130, drivingelement 18, driving element teeth 17, flexible coupling 20, drivenelement 22, and driven element teeth 23. The relationship between thesecomponents and their operation is discussed in detail below inconjunction with FIG. 2, which shows the preferred embodiment of enginecoupling assembly 10.

FIG. 1 represents the prior art wherein a relatively massive flywheel130 forms a part of the coupling assembly 10. Such massive flywheelsused with such coupling assemblies 10 in the prior art have relativelylarge inertias, from approximately 150 to 400 lbs*ft². The primarypurpose of using such a flywheel 130 of high inertia is to smooth outthe amplitude of the torque generated by the flywheel end 117 of enginecrankshaft 116. The large inertias of the prior art flywheels, however,cause the node (location of approximate zero torsional vibrationamplitude) of the first mode of crankshaft torsional vibration, which isequivalent to the second system mode of torsional vibration, to belocated in the vicinity of the flywheel, near the flywheel end 117 ofengine crankshaft 116 in FIG. 1. As a result, the amplitude of thetorsional vibration experienced by the free end 115 of engine crankshaft116 is maximized. The large amplitude of the torsional vibration at thefree end 115 of engine crankshaft 116 places undesirable stress on theteeth of crank pinion gear 118. Furthermore, this stress can betransferred through crank pinion gear 118 to a timing gear train orvehicle accessory drive, thereby shortening the life of thesecomponents.

FIG. 2 shows the preferred embodiment of engine coupling assembly 10 ofthe present invention. Included in engine coupling assembly 10 areflywheel housing 12, flywheel end 117 of engine crankshaft 116, lowinertia flywheel 15, optional ring gear 16, driving element 18, drivingelement teeth 17, retaining ring 19, flexible coupling 20, drivenelement 22, driven element teeth 23, and output shaft 24. In thisconfiguration, output shaft 24 would be a two bearing shaft, i.e. outputshaft 24 would be supported by two bearings (not shown) spaced apartfrom each other and the coupling assembly 10.

As shown in FIG. 2, low inertia flywheel 15 connects to driving element18. Flywheel end 117 of engine crankshaft 116 rigidly connects todriving element 18 and output shaft 24 rigidly connects to drivenelement 22. Driving element 18 and driven element 22 interact throughflexible coupling 20, which is better shown in FIG. 3. Flexible coupling20 is held in place by retaining ring 19 as described in more detailbelow. Optional ring gear 16 may be connected to low inertia flywheel 15to allow for cranking of the engine 110 through driving element 18 andengine crankshaft 116. Flywheel housing 12 preferably encloses both lowinertia flywheel 15 and flexible coupling 20.

In operation, engine crankshaft 116 will supply rotary power to drivingelement 18. Driving element 18 will in turn provide rotary power todriven element 22 and output shaft 24 through flexible coupling 20. Themass of low inertia flywheel 15 is selected so as to locate the node ofthe first mode of crankshaft torsional vibration, which is equivalent tothe second system mode of torsional vibration, approximately in themiddle of engine crankshaft 116, as opposed to having the node locatedon the end of the engine crankshaft 116 near the flywheel 15 as wouldresult from prior art flywheel systems. For example, prior art flywheelsystems used with engines of approximately 500 to 2000 horsepoweremployed flywheels having inertias from approximately 150 to 400lbs*ft². The relatively large inertias associated with these prior artflywheels resulted in a torsional vibration node, having approximatelyzero amplitude, in the engine crankshaft near the junction of thecrankshaft with the flywheel. Furthermore, such large inertias resultedin a torsional vibration having maximum amplitude at the free end 115 ofengine crankshaft 116.

The present invention employs a low inertia flywheel having an inertiaapproximately 5 to 10 times lower, from approximately 15 to 80 lbs*ft²,than that of prior art flywheel systems. Due to the relatively largeinertias associated with the internal components of high horsepowerdiesel engines, applicant has discovered that a high inertia flywheel isnot required, and in fact is detrimental to the operation of engine. Asa result of the use of a low inertia flywheel, the amplitude of thetorsional vibration in engine crankshaft 116 at the free end is reducedby a factor of at least two. This reduction in torsional vibrationamplitude leads to a reduction in stress at the interface between thefree end 115 of engine crankshaft 116 and the drive train connectedthereto, which increases the life expectancy of the drive train.Specifically, by reducing the amplitude at free end 115 of enginecrankshaft 116, the stress on the teeth of crank pinion gear 118 isgreatly reduced. As a result, the life expectancy of crank pinion gear118 is increased. Furthermore, by moving the location of the node of thefirst mode of crankshaft torsional vibration to the center of enginecrankshaft 116, the torsional stress on flywheel end 117 of enginecrankshaft 116 is reduced, thereby increasing the life expectancy ofengine crankshaft 116.

Engine coupling assembly 10 could be an Atra-flex model A-8 couplingassembly manufactured by ATR Inc. The commercially available coupling ismodified for use with the present invention to allow coupling between aflywheel assembly and an output shaft rather than coupling between aninput shaft and output shaft. This modification, however, must be madewithout adding additional mass to the coupling assembly so that thedesired low inertia assembly is achieved.

FIG. 3 shows a cross sectional diagram of the flexible coupling 20 takenalong line 3--3 in FIG. 2. FIG. 3 shows flexible coupling 20, retainingring 19, driving element teeth 17 and driven element teeth 23. As seenin FIG. 3, driving element teeth 17 and driven element teeth 23 areencased in slots 32 of flexible coupling 20. Retaining ring 19circumferentially surrounds flexible coupling 20 to maintain flexiblecoupling 20 in position during operation.

In operation, driving element teeth 17, as a result of the connection tothe engine crankshaft 116 through driving element 18 discussed above inconjunction with FIG. 2, will exert a rotational force on flexiblecoupling 20. This force will be transferred through flexible coupling 20to driven element teeth 23. As a result, power will be transferred tooutput shaft 24 through driven element 22 as discussed above inreference to FIG. 2.

Retaining ring 19 is used to prevent flexible coupling 20 fromdisengaging with driving element teeth 17 and driven element teeth 23.During operation, the centrifugal force exerted on retaining ring 19from flexible coupling 20 will hold retaining ring 19 in position. Whenthe coupling assembly is stationary, however, retaining ring 19 can beeasily removed to allow replacement of flexible coupling 20. Tofacilitate replacement, flexible coupling 20 is provided with a radiallyoriented slit 303, which will allow flexible coupling 20 to be, ineffect, unwound from and easily removed from the coupling and replaced.

Flexible coupling 20 will conform slightly due to the force applied bydriving element teeth 17. By varying the elastomeric properties offlexible coupling 20 (i.e. spring constant), it is possible to fine tunethe resonant frequency of the coupling. The particular advantages andmethods by which this can be done are discussed further below.

Referring now to FIG. 4, an alternate embodiment of the engine couplingassembly 10, which is suitable for use where precise alignment is notneeded and quick changes are desirable, is shown. The structure ofengine coupling assembly 10 is identical to that described above inconnection with FIG. 2 except that retaining ring 19 shown in FIG. 2 hasbeen removed and centering ring 40 has been added. With the addition ofcentering ring 40, retaining ring 19 is no longer needed. The flexiblecoupling 20 is now held in place during operation by centering ring 40.As best seen in FIG. 4, centering ring 40 is rigidly connected to drivenelement 22 and surrounds flexible coupling 20. By using centering ring40 in this way, the concentricity of output shaft 24 relative to drivingelement 18 is maintained. Therefore, in this configuration, only onebearing is required to be used with output shaft 24, which acts as afloating shaft. Output shaft 24 could be connected to a flexplateassembly 401 or gear coupling (not shown). This coupling configurationis particularly advantageous when frequent quick changes are desirableand precise alignment is not necessary. Specific uses could includedynamometer drives and other engine or load test applications.

A third embodiment of the engine coupling assembly 10 is shown in FIG.5. Here, an integral bearing assembly is incorporated into enginecoupling assembly 10. The bearing assembly contains support flanges 503,bearings 505, and driven element 22. Support flanges 503 are rigidlyconnected to flywheel housing 12 and support driven element 22 throughbearings 505. The bearing assembly rotatably supports output shaft 24,which obviates the need for a separate bearing assembly to beconstructed somewhere along the length of the driven shaft.Additionally, the bearing assembly ensures that output shaft 24 will bemaintained in a concentric relationship with engine coupling assembly10. A bearing assembly of this type would typically be used to drive acardan type shaft or single bearing generator. However, other possibleshaft configurations could be used where an integral bearing assembly isadvantageous.

FIG. 6 is a Campbell, or interference, diagram showing the relationshipbetween system mode frequencies, engine RPM, and engine excitationorders, which represent the number of vibrations occurring perrevolution of the engine crankshaft 116. The Y-axis of the graphrepresents the natural frequency of a system mode in hertz. The X-axisof the graph represents the engine speed in rotations-per-minute (RPM).The engine excitation orders of 0.5, 1.5, 2, 2.5, 3, 3.5, 4, 4.5, and 6are plotted as lines on the graph.

The natural resonant frequency of the first system mode, or couplingmode, of torsional vibration will be determined by the properties of theelastomeric coupling used in the engine coupling assembly. Inparticular, the natural frequency, f, of the first system mode oftorsional vibration is given by equation 1 below, ##EQU1## where K isthe spring constant of the elastomeric coupling and I_(e) is theeffective inertia of the system given by equation 2, ##EQU2## where I₁represents the inertia of the engine components, flywheel, and couplingcomponents attached thereto, and I₂ represents the inertia of the drivencoupling components, driven shaft, and driven load.

Solving equation 1 and equation 2 with respect to the parametersassociated with prior art flywheel systems yields a natural resonantfrequency of the first system mode of torsional vibration ofapproximately 15-20 hertz. If a typical engine operating range ofapproximately 600 to 2400 RPM is used, FIG. 6 shows that the 0.5 engineexcitation order will be present at a 15-20 hertz natural frequency fromapproximately 1800-2400 RPM.

The 0.5 order is typically not harmful if the engine is operatingsmoothly. The 0.5 order can, however, be excited as a result of amisfiring cylinder, by the action of an engine speed governor, or fromany other occurrence that results in erratic engine operation. If thisoccurs, the resulting resonance will quickly destroy the elastomericcoupling and require expensive and time consuming repairs.

By referring to FIG. 6, it can be seen that if the first system moderesonant frequency can be raised above approximately 20 hertz, then the0.5 order mode will no longer be present across the engine operatingrange from 600 to 2400 RPM. By referring to equation 1, it can be seenthat the first system mode resonant frequency, f, can be raised ineither of two ways: (1) by raising the spring constant, K, of thecoupling, or (2) by lowering the effective inertia, I_(e), of thesystem.

The first method, raising the spring constant, K, of the coupling willonly result in minor fluctuations in the resonant frequency. This is sobecause the spring constant must be retained within certain limits inorder to maintain the benefits of the flexible coupling. If the springconstant is raised to a point to advantageously effect the resonantfrequency, the coupling will no longer be elastomeric. The elasticity ofthe coupling is required and for this reason, the natural resonantfrequency of the coupling cannot be significantly altered by changingthe spring constant of the coupling.

The second method of raising the first system mode resonant frequency,by lowering the effective inertia of the system, can be usedeffectively. From equation 2, it can be seen that the effective inertiacan be lowered by lowering either I₁ or I₂. I₂, however, is determinedby the load components and is usually beyond the control of the enginemanufacturer. Therefore, it is necessary to lower I₁, the inertia of theengine components, flywheel, and coupling components attached thereto.

The present invention recognizes this deficiency in prior art systemsand solves the problem by providing a low inertia flywheel and couplingassembly. In the system of the present invention, the first system moderesonant frequency is raised to approximately 22-25 hertz, therebyavoiding the 0.5 order mode over the entire operating range of theengine. This is accomplished by selecting a mass for the flywheelassembly such that the inertia of the system results in a frequency, f,of between approximately 22 and 25 hertz as determined from equations 1and 2 above. Furthermore, the present invention uses the first methoddiscussed above, that of raising the spring constant of the elastomericcoupling, to fine tune the resonant frequency. Slight changes in thespring constant can result in minor alterations to the resonantfrequency without effecting the advantages of the flexible coupling,thereby allowing a fine tuning of the coupling resonant frequency. Withthe improved coupling assembly, the 0.5 order is eliminated across theentire normal operating range of the engine. Therefore, coupling failureas a result of erratic engine operation exciting the 0.5 order iscompletely avoided with the present invention.

A third advantage of the present invention can be seen by reference toFIGS. 6 and 7. By reducing the mass of flywheel 15, the primary engineorders are changed from whole number orders to fractional orders,preferably odd multiples of 0.5, i.e. 0.5, 1.5, 2.5, 3.5, etc. Theprimary engine vibration orders include second, third, fourth and sixthorders for a four, six, eight and twelve throw crank respectively. Thechange in the primary engine vibration orders is particularlyadvantageous because it causes the stress at the free end 115 of enginecrankshaft 116 to be distributed over at least twice as many gear teethof crank pinion gear 118. In conjunction with the reduction in torsionalvibration amplitude that results from a crank center node, this reducedthe stress on the gear teeth of crank pinion gear 118 by at least fourtimes.

For example, referring to FIG. 6, if an engine is operating at asynchronous speed of 1800 RPM, and the first mode of crankshafttorsional vibration (second system mode of vibration) has a naturalfrequency of 120 hertz, the engine will experience fourth ordervibrations. This results in exactly four vibration pulses, spaced 90degrees apart, per revolution of engine crankshaft 116. Each vibrationpulse will occur at exactly the same location (0, 90, 180 and 270degrees) during each revolution. This causes undesirable vibrationalstresses to be concentrated on the same gear teeth during eachrevolution, which leads to earlier failure of these gear teeth. Thepresent invention, however, by reducing the mass of engine flywheel 15,will cause the engine orders to be changed from whole orders to halforders. As a result, the fourth order of the above example changes to3.5 and 4.5. This results in 3.5 and 4.5, respectively, vibration pulsesper revolution of engine crankshaft 116. It will now require twocomplete revolutions before the vibration pulses will occur on the samegear teeth, thereby distributing the stress over twice as many gearteeth. With a 3.5 primary engine excitation order, the vibration pulseswill occur approximately every 102.9 degrees (3.5 pulses per 360 degreerevolution equals 1 pulse every 102.9 degrees). With a 4.5 primaryengine excitation order, the vibration pulses will occur every 80degrees (4.5 pulses per 360 degree revolution equals 1 pulse every 80degrees). The resulting distribution of stress over twice as many gearteeth will result in much longer component life.

The 0.5 order change that results from the present invention can bebetter seen in FIG. 7. FIG. 7 depicts the resulting engine orders for aprior art near flywheel node and for the crank center node resultingfrom the present invention for four, six and eight throw crankshafts.For example, an eight throw crankshaft employing a prior art flywheelsystem in which a torsional vibration node occurs at or near theflywheel results in prominent fourth and eighth orders of torsionalvibration. As a result, the stress on a crank pinion gear will beconcentrated on the same four or eight teeth respectively. However, if alow inertia flywheel is employed, that results in a crank center node oftorsional vibration, FIG. 7 shows that the prominent engine ordersbecome 2.5, 5.5 and 6.5. The resulting stress on a crank pinion gearwill therefore be distributed over 5, 10, or 12 teeth respectively.

The following table illustrates the above effect:

                  TABLE I                                                         ______________________________________                                                     Natural  Vibrational Torque                                      Inertia      Resonant By Engine Order                                         (In-Lb-      Frequency                                                                              (In-Lbs.)                                               Coupling                                                                              Sec.sup.2)                                                                             (Hertz)  3.5    4.0    4.5                                   ______________________________________                                        Heavy   105      14.5      6825  25611   8285                                 DCB 834.5                                                                     Prior Art                                                                     Light    18      24.6     10394   7218  22320                                 Atra A-8                                                                      Present                                                                       Invention                                                                     ______________________________________                                    

Table I illustrates the changes that occur in inertia, natural resonantfrequency, and vibrational torque for 3.5, 4, and 4.5 engine excitationorders when a low inertia flywheel of the present invention is used inplace of a high inertia flywheel of a typical prior art system. As seenin Table I, the inertia of the system of the present invention is 18inch-pound-second² ; almost a factor of 6 times smaller than the 105inch-pound-second² inertia of a typical prior art system. Furthermore,Table I shows that the natural resonant frequency of the improvedcoupling assembly increases from 14.5 Hertz to 24.6 Hertz.

The change of the primary engine excitation orders from whole orders tohalf orders is clearly illustrated by Table I. With a high inertiacoupling, as used in the prior art, the torque due to the fourth orderis 25,611 inch-pounds, while that of the 3.5 and 4.5 orders is 6,825 and8,285 inch-pounds respectively. With the low inertia flywheel of thepresent invention, the respective torque for the 3.5, 4, and 4.5 ordersis 10,394, 7,218, and 22,320 inch-pounds. As described above, Table Ishows that the fourth excitation order has reduced by approximately18,000 inch-pounds, while the 3.5 order has increased by approximately3,500 inch-pounds and the 4.5 order has increased by approximately14,000 inch-pounds. Table I also shows that the peak torque of 25,611inch-pounds present in the fourth order of a prior art flywheel systemis reduced to 22,320 inch-pounds in the present invention. Therefore,not only is the torque spread out over more teeth of a driven gear, butthe peak torque value is reduced.

Raising the natural resonant frequency of the coupling mode (firstsystem mode) is also particularly advantageous during engine start-upand shut down. During engine start-up, the rotational speed of theengine will increase from zero to the idle speed. When this occurs, theengine will be required to pass through the primary order of torsionalvibration at the resonant frequency of the coupling assembly. Ideally,it is desirable to have the engine RPM associated with the primaryexcitation order and coupling resonant frequency to be close, within 50RPM, but below the engine low-idle speed. This is so because themagnitude of the deflections resulting from the primary order areinversely related to the square of the engine RPM.

For example, in an eight throw crank engine the primary exciting orderis the fourth order. If the low-idle speed of the engine is 400 RPM,then it is desirable to have the engine RPM associated with the couplingresonant frequency and the fourth excitation order to be betweenapproximately 350 and 400 RPM, and preferably between 390 and 400 RPM.From FIG. 6, it can be seen that at a coupling resonant frequencytypical of prior art systems of 16 hertz, the engine speed for a fourthorder excitation is approximately 240 RPM. If the coupling resonantfrequency is increased to 25 hertz, the engine speed for a fourth orderexcitation increases to approximately 380 RPM. Therefore, the magnitudeof the deflection, which is proportional to the square of the inverse ofthis difference in RPM will be approximately 2.5 times less in a systemwith a coupling resonant mode frequency of 25 hertz than in a systemwith a resonant frequency of 16 hertz. This reduction in magnitude ofthe deflection will in turn result in a reduction in the stresses thatoccur on the coupling assembly during engine start-up.

I claim:
 1. A method of reducing stress between one end of a rotatingcrankshaft of an internal combustion engine and a drive train wherein acoupling assembly including a flywheel means transmits rotary powerbetween a second end of said crankshaft opposite the drive train and adriven shaft, the mass of said flywheel being such that the node of thefirst crankshaft mode is located in the vicinity of the flywheel means,consisting of the single step of reducing the mass of the flywheel meanssuch that said node is moved to the vicinity of the middle of saidcrankshaft to reduce the amplitude of the torsional deflection of thefirst end of the crankshaft.
 2. The method of claim 1, wherein saidflywheel mass is reduced by replacing it with a second flywheel meanshaving a lower mass than the original flywheel means.
 3. The stressreducing method defined in claim 1, wherein said coupling assemblyincludes a flexible coupling, and the reduction of the mass of theflywheel increases the natural frequency of the coupling assembly to alevel higher than said 0.5 order of torsional vibration.
 4. An improvedcoupling assembly as defined in claim 3, wherein said flexible couplingincludes an elastomeric element characterized by a specific hardness,and further comprising the step of adjusting the natural frequency ofsaid coupling assembly by adjusting the hardness of said elastomericelement.
 5. An improved coupling assembly as defined in claim 4, whereinsaid hardness of said elastomeric element is adjusted by replacing itwith another elastomeric element having a different hardness.